In sliding vane positive displacement pumps, such pumps are used in a number of different industrial and commercial processes to force fluid movement from a first location to a second location. One example of a sliding vane pump of this type is illustrated in FIGS. 1 and 2.
The prior art sliding vane pump 10 includes a housing or casing 11 that defines a hollow section which is shaped to define a pump chamber 12. Typically, the pump chamber 12 is defined by a liner 13 that is stationarily supported in the casing 11 and has an eccentric, non-circular cross-sectional profile. The pump chamber 12 is supplied with process fluid through an inlet 15 and discharges from an outlet 16, which inlet 15 and outlet 16 respectively open into and out of the pump chamber 12.
In prior art pumps 10 of this type, flat, stationary discs 17 and 18 define the front and rear ends of the chamber 12. The discs 17 and 18 are stationary and are confined axially between a first head 21 and a second head 22 which generally enclose the front and rear ends of the pump chamber 12. The first and second heads 21 and 22 are affixed to the casing 11 by fasteners and sandwich the discs 17 and 18 and the liner 13 therebetween so as to prevent movement of these components during shaft rotation.
A shaft 24 extends through the casing 11 and has an inboard first end 25, which projects from the casing 11 and is driven by a motor or other motive means, and an outboard second end 26. In this design, the second shaft end 26 terminates within the casing 11 and is rotatably supported by the outboard head 22. The shaft ends 25 and 26 are supported by bearings 27 and 28 which are respectively supported within corresponding channels in the heads 21 and 22 and rotatably support the shaft 24 to permit rotation thereof. The bearings 27 and 28 are retained axially in position by bearing locknuts 30 and 31, which thread onto the shaft ends 25 and 26, and in turn, are enclosed by bearing covers 32 and 33, which are removably affixed to the heads 21 and 22.
The shaft 24 extends through the pump chamber 12 by extending axially through shaft holes 35 and 36 which are formed in the center of the discs 17 and 18. A small radial gap is defined between the inside diameter of the shaft holes 35 and 36 and the opposing outside shaft surface 37, and while some process fluid might leak axially out of the pump chamber 12 along the radial gaps, mechanical seals 40 and 41 are provided which seal radially between the casing 11 and shaft 24 to prevent leakage of such fluid out of the pump 10.
To effect pumping, attached to the shaft 24 is a rotor 45 that is secured to the shaft 24 so as to rotate in unison therewith. The rotor 45 is located within the pump chamber 12 to draw fluid through the inlet 15 and discharge process fluid through the outlet 16. The rotor 45 includes vane slots 46 which are spaced circumferentially from each other. These vane slots 46 open radially outwardly, and also open axially through the opposite rotor faces 45A.
Normally, vanes (not shown in FIGS. 1 and 2) project outwardly from the slots 46 in the rotor 45, although the vanes are movable radially into and out of the slots 46. The vanes are confined axially within the slots 46 by the stationary discs 17 and 18 which are positioned axially adjacent to the rotor 45. As the shaft 24 and rotor 45 turn, the volume of the space in the chamber 12 between circumferentially adjacent vanes and the radially opposed surfaces of the rotor 45 and liner 13 (each space referred to as a fluid cavity), cyclically increases and decreases due to the eccentric profile defined by the liner 13. As a result of the increase in volume of a fluid cavity as it begins to travel away from the inlet 15, a suction is formed in the cavity. The suction draws fluid into the fluid cavity through the inlet 15. As the rotor continues to turn, owing to the geometry of the pump chamber 12 and liner 13, the volume of the fluid cavity decreases as it travels towards the outlet 16. As a result of the volume of the cavity decreasing, the fluid in the cavity is discharged through an outlet 16.
In the known configuration, the liner 13 and discs 17 and 18 remain stationary while the rotor 45 rotates relative thereto. The discs 17 and 18 are located at the opposite ends of the rotor 45 and respectively include disc faces 17A and 18A which face axially toward the opposing rotor faces 45A. Due to the relative rotation therebetween, a small axial clearance or end clearance is required between the disc faces 17A and 18A and the rotor faces 45A. Typically, the discs 17 and 18 and the rotor 45 are metallic, and as such, contact must be avoided during shaft rotation, wherein such face contact can cause galling between these components. In these pump designs, it thereby may be desired to provide expensive coatings on the heads and discs 17 and 18 to prevent galling damage.
Due to this end clearance, however, disadvantages are present with known pump designs. More particularly, the opposed end faces 17A, 18A and 45A and the end clearances therebetween generate dynamic sealing due to the relative movement of the rotor end faces 45A. As a result, the dynamic movement of the components impedes leakage of fluid between such end faces 17A, 18A and 45A. However, these end clearances still define paths that extend facewise across the end faces 45A and that allow pressurized fluid to slip from the outlet side to the inlet side of the rotor 45 which thereby reduces the overall hydraulic efficiency of the pump 10, since such fluid is not discharged through the outlet 16 but instead returns to the inlet side and is then displaced again by the rotor 45 and vanes back towards the outlet 16. This loss is conventionally known as slip.
While it is desirable to minimize the end clearance to minimize slip, this minimizing of the axial clearance space results in tight dimensional tolerances for the pump components and requires precise positioning of the rotor 45 between the two discs 17 and 18. In one negative aspect of this known design, the axial location of the rotor 45 and discs 17 and 18 must be precise.
In a second aspect, the rotor 45 has a much larger diameter than the shaft 24 and the rotor faces 45A and disc faces 17A and 18A extend radially a significant dimension. In other words, the outside diameters (OD) of the rotor 45 and discs 17 and 18 are spaced radially outwardly of the shaft by a significant distance, such that the rotor faces 45A and disc faces 17A and 18A have a significant radial width as measured radially outwardly from the shaft 24 to the OD of each disc 17/18 and rotor 45. To maintain a constant and uniform axial clearance facewise across this radial width, it also is important that the opposed faces 17A and 18A be parallel to each other and perpendicular to the shaft axis. The large diameter of the rotor 45 relative to the shaft 24 creates a need for a tight or precise perpendicularity and machining tolerances between the rotor 45 and shaft 24 and between the heads 21 and 22 and respective discs 17 and 18.
Even if the end clearances are minimized, the overall area or radial width of the end clearances is still relatively large and this defines significant area over which slip can occur. Hence, these pump designs still exhibit disadvantages resulting from the slip which occurs between the stationary pump components and the rotor 45.
In other pump designs as disclosed in U.S. Pat. No. 7,134,551 (Bohr) and U.S. Pat. No. 7,316,551 (Bohr), these designs relate to variations of a rotary vane, positive displacement pump. One such pump embodying this invention has a rotor that is attached to the front end of the complementary shaft. An inboard disc is located between the rotor and shaft to form a first end surface against which the pump vanes seat. In another such pump, a second disc may be fitted over the opposed front end of the rotor to form the second end surface against which the vanes seat.
In another such pump, a second rotor may be fixed with respect to the opposed front face of the second disc. In another such pump, separate pump chambers are provided for corresponding rotors. In another such pump, a third disc may be fitted over the opposed front end of the second rotor. The discs rotate in unison with the rotor(s) and the shaft. These designs do not have a bearing supported forward end.
In these pump designs, the discs extend radially beyond the outside rotor diameter and as such, the discs have disc faces which face towards the side faces of a liner. The discs rotate relative to the liner and define end faces which face axially toward liner end faces. These opposed faces are relatively movable, and create clearance spaces that can permit slip therebetween. Further, the axial positioning of the discs and liner must be maintained precisely. Here again, it is desirable to provide a pump design which provides improved performance over these known pump designs.